1. Field of the Invention
The present invention relates to a turbulence generating method making effective use of a swirl of intake air to efficiently generate turbulence to thereby improve the combustion efficiency of a reciprocating internal combustion engine, and an engine for effecting this method.
2. Description of the Prior Art
In order to improve combustion in a reciprocating piston internal combustion engine to thereby improve fuel economy and the output power, and to reduce the noxious content of the exhaust gas, it is most important to increase the flow velocity of the gas in the combustion chamber to thereby augment turbulence. If this is satisfied, the burning velocity can be increased to shorten the time period for combustion.
In order to augment fluid turbulences with a view to achieving this object, according to the prior art, the following three methods have been conceived and carried out:
(1) intensifying the squish; PA1 (2) intensifying the intake swirl; and PA1 (3) combining the intake swirl and the squish.
In the so-called "premixed type internal combustion engine" such as a spark ignition type internal combustion engine for injecting fuel into an intake port or an intake pipe or a spark ignition type internal combustion engine using a carburetor, in order to raise the compression ratio to improve the fuel economy, it is necessary to intensify the turbulence in order to raise the flame propagation velocity to thereby shorten the combustion period, so that combustion may be completed before knocking occurs.
For this purpose, according to the first method, a combustion chamber 3 which is defined by the top face of a piston 1 and a cylinder head 2 is formed at a portion with a recess 4, as shown in FIG. 1, so that an intense squish (as indicated by arrows) may be generated immediately before top dead center as the piston 1 rises.
According to the second method, there is provided a mechanism for generating a swirl in the intake air-fuel mixture, e.g., a shrouded intake valve, a tangential intake port or a helical port, to generate a swirl during the intake stroke so that turbulence may be intensified by the swirl remaining during the compression stroke.
The third method combines the preceding two methods.
The relationship between heat release and crank angle in each method, as examined by use of a pressure indicator attached to the combustion chamber to measure pressure change, is illustrated in FIG. 2.
Although the second method making use of the swirl provides faster combustion (the combustion is faster for larger gradients in FIG. 2) than any method using neither swirl nor squish, the turbulences generated during the suction stroke are soon attenuated when using only the swirl generated during the intake stroke, and only a swirl is left to provide a weak combustion promoting effect.
Next, according to the first method making use of squish, the squish flow velocity is maximum about 10 degrees before the top dead center (TDC) position of the piston, and the squish itself is then abruptly attenuated so that the combustion is relatively fast from the initial stage to the intermediate stage of combustion, as shown in FIG. 2. However, the turbulence generation disappears with the attenuation of the squish so that the latter half of the combustion is not fast. Since the third method employs a recess in a portion of the piston top face to use both the squish flow and the intake swirl, as shown in FIG. 4, the swirl in the recess is intensified to the extent corresponding to the difference in diameter between the cylinder and the recess in the vicinity of top dead center to effect the fastest combustion in FIG. 2 by combining the swirl and squish effects. As has been described above, however, the squish flow is quickly attenuated so as to exhibit no combustion promoting effect at the latter half stage of combustion.
As has been described above, the swirl has no considerable effect, but the squish or its combination with the swirl accelerates the burning velocity. However, this combustion promoting effect is obtained only at the initial stages of combustion. In order to intensify the squish, on the other hand, the clearance between the piston and the lower surface of the cylinder head at TDC has to be no more than 1 mm, so that the heat loss at that portion is increased and the thermal efficiency is decreased.
Thus, the burning velocity cannot be accelerated over the whole range of the combustion period by any of the above methods. There also arises a problem in that the compression ratio cannot be made sufficiently high without inviting knocking. At the present technical level, the compression ratio is limited to 9 to 9.5 at the highest for an engine cylinder diameter of 80 to 90 mm and a fuel octane number of 90.
On the other hand, the aforementioned third method is used in either a direct injection stratified charge engine which is equipped with a fuel injection valve in the combustion chamber and with ingniting means such as a spark plug, or in a direct injection type Diesel engine which is equipped with a fuel injection nozzle in the combustion chamber, and in which the compression ratio is raised until spontaneous ignition.
In the former engine, the fuel is injected during the intake stroke and to about 60 degrees before TDC to provide the homogeneous combustion whereas, in the latter engine, the fuel is injected during the period 60 degrees before and at TDC. Since the fuel is injected directly into the combustion chamber in either case, the intake air is cooled down by the latent heat of evaporation, or combustion is started at the instant when the air and the fuel are mixed so that a high compression ratio can be employed without the danger of knocking, unlike the spark ignition type internal combustion engine using a carburetor.
However, the combustion chambers of these engines are formed with a recess in the piston and with an intake port for generating a swirl during the intake stroke to effect turbulence generation by the aforementioned third method. As has been described hereinbefore, intense turbulences are generated immediately before TDC to promote the mixing of the air and the fuel. However, the turbulence intensity is too quickly attenuated to promote combustion after TDC, i.e., the intermediate stage of combustion, to fail to accelerate combustion at the latter half thereof, as shown in FIG. 2, so that even a rise in the compression ratio will not lead to a significant improvement in the fuel economy. Since the fuel is injected directly into the combustion chamber, moreover, the hydrocarbons, i.e., HC left unevaporated, are liable to be emitted. In the direct injection type Diesel engine, on the other hand, the fuel injection at maximum is started at 20 to 10 degrees before TDC and is ended at TDC to about 10 degrees after TDC, whereas combustion is started at 10 to 5 degrees before TDC and is ended at 40 to 50 degrees after TDC. In other words, the latter half of the fuel injection is conducted after combustion has already been started. Since the compression ignition type direct injection (i.e., Diesel) engine also uses the third method in which both the swirl generated during the suction stroke and the squish generated by forming the piston with a recess (e.g., of shallow or deep dish shape or of troidal or ball-in shape) are used, it is advantageous for the purpose of the dispersion of the fuel at the preparation of the mixture, especially, the initial mixture preparation. For mixture preparation around or after TDC or after the intermediate stage of combustion, however, the turbulence intensity is insufficient to cause a deterioration of fuel economy or the generation of smoke.